Turbochargers are a type of forced induction system. They deliver air, at greater density than would be possible in the normally aspirated configuration, to the engine intake, allowing more fuel to be combusted, thus boosting the engine's horsepower without significantly increasing engine weight. This can enable the use of a smaller turbocharged engine, replacing a normally aspirated engine of a larger physical size, thus reducing the mass and aerodynamic frontal area of the vehicle.
Turbochargers (FIG. 1) use the exhaust flow (100), which enters the turbine housing at the turbine inlet (51) of the turbine housing (2), from the engine exhaust manifold to drive a turbine wheel (70), which is located in the turbine housing. The turbine wheel is solidly affixed to a shaft, the other end of which contains a compressor wheel which is mounted to the shaft and held in position by the clamp load from a compressor nut. The primary function of the turbine wheel is providing rotational power to drive the compressor. Once the exhaust gas has passed through the turbine wheel (70) and the turbine wheel has extracted energy from the exhaust gas, the spent exhaust gas (101) exits the turbine housing (2) through the exducer (52) and is ducted to the vehicle downpipe and usually to the after-treatment devices such as catalytic converters, particulate and NOx traps.
The power developed by the turbine stage is a function of the expansion ratio across the turbine stage. That is the expansion ratio from the turbine inlet (51) to the turbine exducer (52). The range of the turbine power is a function of, among other parameters, the flow through the turbine stage.
The compressor stage consists of a wheel and its housing. Filtered air is drawn axially into the inlet (11) of the compressor cover (10) by the rotation of the compressor wheel (20). The power generated by the turbine stage to the shaft and wheel drives the compressor wheel (20) to produce a combination of static pressure with some residual kinetic energy and heat. The pressurized gas exits the compressor cover (10) through the compressor discharge (12) and is delivered, usually via an intercooler to the engine intake.
The design of the turbine stage is a compromise among the power required to drive the compressor; the aerodynamic design of the stage; the inertia of the rotating assembly, of which the turbine is a large part since the turbine wheel is manufactured typically in Inconel which has a density 3 times that of the aluminum of the compressor wheel; the turbocharger operating cycle which affects the structural and material aspects of the design; and the near field both upstream and downstream of the turbine wheel with respect to blade excitation.
Part of the physical design of the turbine housing is a volute, the function of which is to control the inlet conditions to the turbine wheel such that the inlet flow conditions provide the most efficient transfer of power from the energy in the exhaust gas to the power developed by the turbine wheel. Theoretically the incoming exhaust flow from the engine is delivered in a uniform manner from the volute to a vortex centered on the turbine wheel axis. To do this, the cross sectional area of the volute gradually and continuously decreases until it becomes zero. The inner boundary of the volute can be a perfect circle, defined as the base circle; or, in certain cases, such as a twin volute, it can describe a spiral, of minimum diameter not less than 106% of the turbine wheel diameter. The volute is defined by the decreasing radius of the outer boundary of the volute and by the inner boundary as described above, in one plane defined in the “X-Y” axis as depicted in FIG. 4, and the cross sectional areas, at each station, in the plane passing through the “Z” axis, as depicted in FIG. 16. The “Z” axis is perpendicular to the plane defined by the “X-Y” axis and is also the axis of the turbine wheel.
The design development of the volute initiates at slice “A”, which is defined as the datum for the volute. The datum is defined as the slice at an angle of “P” degrees above the “X-axis of the turbine housing containing the “X”-axis, “Y”-axis and “Z”-axis details of the volute shape.
The size and shape of the volute is defined in the following manner: The widely used term A/R represents the ratio of the partial area at slice “A” divided by the distance from the centroid (161) of the shaded flow area (160) to the turbo centerline. In FIGS. 15A and 15B the centroids (161) determine the distance RA and RB to the turbo centerline. For different members of a family of turbine housings, the general shape remains the same, but the area at slice “A” is different as is the distance RA. The A/R ratio is generally used as the “name” for a specific turbine housing to differentiate that turbine housing from others in the same family (with different A/R ratios). In FIG. 15A. the volute is that of a reasonably circular shape. In FIG. 15B the volute is that of a divided turbine housing which forces the shape to be reasonably triangular. Although the areas at slice “A” for both volutes are the same, the shapes are different and the radii to the centroids are different (due to the volute shape), so the A/Rs will be different. Slice “A” is offset by angle “P” from the “X”-axis. The turbine housing is then geometrically split into equal radial slices (often 30°, thus at [30x+P]°, and the areas (AA-M) and the radii (RA-M) along with other geometric definitions such as corner radii are defined. From this definition, splines of points along the volute walls are generated thus defining the full shape of the volute. The wall thickness is added to the internal volute shape and through this method a turbine housing is defined.
The theoretically optimized volute shape for a given area is that of a circular cross-section since it has the minimum surface area which minimizes the fluid frictional losses. The volute, however, does not act on its own but is part of a system; so the requirements of flow in the planes from slice “A”, shown in FIG. 4 to the plane at slice “M”, and from “M” to the tongue, influence the performance of the turbine stage. These requirements often result in compromises such as architectural requirements outside of the turbine housing, method of location and mounting of the turbine housing to the bearing housing, and the transition from slice “A” to the turbine foot (51) result in turbine housing volutes of rectangular or triangular section, as well as in circular, or combinations of all shapes. The rectangular shape of the volute (53) in FIG. 1, showing a section “D-K” is a result of the requirement not only to fit VTG vanes into the space such that the flow is optimized through the vanes and that the vanes can be moved and controlled by devices external to the turbine housing, but also to minimize the outline of the turbine housing so the turbocharger fits on an engine.
The turbine housing foot is usually of a standard design as it mates to exhaust manifolds of many engines. The foot can be located at any angle to, or position relative to, the “volute”. The transition from the foot gas passages to the volute is executed in a manner which provides the best aerodynamic and mechanical compromise.
The roughly triangular shape of the volute in FIG. 2, taken at the same sections as those above, is the more typical volute geometry for fixed and wastegated turbine housings. The addition of the divider wall (21) is to reduce aerodynamic “cross-talk” between the volutes in an effort to maintain pulse flow, from a divided manifold, to harvest the pulse energy in the work extracted by the turbine wheel. The pressure pulses in the exhaust manifold are a function of the firing order of the engine.
Turbine housings are typically designed in families (typically up to 5 in a family) which use turbine wheels of the same diameter, or a group of wheels with close to the same diameter. They may use the same turbine foot size. For example, a family of turbine housings for a 63 mm turbine wheel may cover a range of A/Rs from 1.8 to 2.2. FIG. 5 depicts the area schedule for three volutes of a family. The largest volute is a 1.2 A/R volute, shown by the dotted line (40). The smallest volute is a 0.8 A/R volute; shown by the dashed line (41) and the mean volute, in the middle of the family, is shown by the solid line. The X-axis depicts the angle of the slice, from 30° (section “A”) to 360° (the tongue); the Y-axis depicts the area of the section at the respective angle.
Some turbine wheels are specifically designed to harness this pulse energy and convert it to rotational velocity. Thus the conversion of pressure and velocity from the exhaust gas for a pulse flow turbine wheel in a divided turbine housing is greater than the conversion of pressure and velocity from a steady state exhaust flow to the turbine wheel velocity. This pulse energy is more predominant in commercial Diesel engines, which operate at around 2200 RPM, with peak torque at 1200 to 1400 RPM, than in gasoline engines which operate at much higher rotational speed, often up to 6000 RPM, with peak torque at 4000 RPM so the pulse is not as well defined.
The basic turbocharger configuration is that of a fixed turbine housing. In this configuration the shape and volume of the turbine housing volute (53) (FIG. 1) is determined at the design stage and cast in place.
The next level of sophistication is that of a wastegated turbine housing. In this configuration the volute is cast in place, as in the fixed configuration above. In FIG. 2, the wastegated turbine housing features a port (54) which fluidly connects the turbine housing volute (53) to the turbine housing exducer (52). Since the port on the volute side is upstream of the turbine wheel (70), and the other side of the port, on the exducer side, is downstream of the turbine wheel, flow through the duct connecting these ports bypasses the turbine wheel (70), thus not contributing to the power delivered to the turbine wheel.
The wastegate in its most simple form is a valve (55), which can be a poppet valve. It can be a swing type valve similar to the valve in FIG. 2. Typically these valves are operated by a “dumb” actuator which senses boost pressure or vacuum to activate a diaphragm, connected to the valve, and operates without specific communication to the engine ECU. The function of the wastegate valve, in this manner, is to cut the top off the full load boost curve, thus limiting the boost level to the engine. The wastegate configuration has no effect on the characteristics of the boost curve until the valve opens. More sophisticated wastegate valves may sense barometric pressure or have electronic over-ride or control, but they all have no effect on the boost curve until they actuate to open or close the valve.
FIGS. 6A and 6B represent compressor maps. The “Y” axis (61) represents the boost or pressure ratio level and the “X” axis (60) represents the expansion ratio. FIG. 6A depicts the boost curve (67) for a fixed turbine housing. In this configuration as the turbo speed rises the upper part (65) of the boost curve continues to increase in pressure ratio as the mass flow through the wheel continues to increase. FIG. 6B depicts the boost curve (68) for a wastegated turbine housing of the same A/R as that for FIG. 6A, or a wastegated turbine housing in which the wastegate valve did not open. In FIG. 6B it can be seen that the lower shape of the boost curve (68) is exactly the same as the lower part boost curve (67) in FIG. 6A to the point (66) at which the valve opens. After this point, the boost curve (62) is relatively flat, so as the turbo speed increases the boost curve is controlled at a max. level while the massflow through the wheel continues to increase. While a wastegate can be used to limit boost levels, its turbine power control characteristics are rudimentary and coarse.
A positive byproduct of wastegated turbine housings is the opportunity to reduce the A/R of the turbine housings. Since the upper limit of the boost is controlled by the wastegate, a reduction in A/R can provide better transient response characteristics. If the wastegated turbocharger has a “dumb” actuator, which operates on a pressure or vacuum signal only, and is operated at altitude, then the critical pressure ratio at which the valve opens is detrimentally affected. Since the diaphragm in the actuator senses boost pressure on one side, and barometric pressure on the other, the tendency is for the actuator to open later (since the barometric pressure at altitude is lower than that at sea level) resulting in over-boost of the engine.
Engine boost requirements are the predominant drivers of compressor stage selection. The selection and design of the compressor is a compromise between the boost pressure requirement of the engine; the mass flow required by the engine; the efficiency required by the application; the map width required by the engine and application; the altitude and duty cycle to which the engine is to be subjected; the cylinder pressure limits of the engine; etc.
The reason this is important to turbocharger operation is that the addition of a wastegate to the turbine stage allows matching to the low speed range with a smaller turbine wheel and housing. Thus the addition of a wastegate brings with it the option for a reduction in inertia. Since a reduction in inertia of the rotating assembly typically results in a reduction of particulate matter (PM), wastegates have become common in on-highway vehicles. The problem is that most wastegates are somewhat binary in their operation, which does not fit well with the linear relationship between engine output and engine speed.
The next level of sophistication in boost control of turbochargers is the VTG (the general term for variable turbine geometry). Some of these turbochargers have rotating vanes; some have sliding sections or rings. Some titles for these devices are: Variable turbine geometry (VTG), Variable geometry turbine (VGT), variable nozzle turbine (VNT), or simply variable geometry (VG).
VTG turbochargers utilize adjustable guide vanes FIGS. 3A and 3B, rotatably connected to a pair of vane rings and/or the nozzle wall. These vanes are adjusted to control the exhaust gas backpressure and the turbocharger speed by modulating the exhaust gas flow to the turbine wheel. In FIG. 3A the vanes (31) are in the minimum open position. In FIG. 3B the vanes (31) are in the maximum open position. The vanes can be rotatably driven by fingers engaged in a unison ring, which can be located above the upper vane ring. For the sake of clarity, these details have been omitted from the drawings. VTG turbochargers have a large number of very expensive alloy components which must be assembled and positioned in the turbine housing so that the guide vanes remain properly positioned with respect to the exhaust supply flow channel and the turbine wheel over the range of thermal operating conditions to which they are exposed. The temperature and corrosive conditions force the use of exotic alloys in all internal components. These are very expensive to procure, machine, and weld (where required). Since the VTG design can change turbocharger speed very quickly, extensive software and controls are a necessity to prevent unwanted speed excursions. This translates to expensive actuators. While VTGs of various types and configurations have been adopted widely to control both turbocharger boost levels and turbine backpressure levels, the cost of the hardware and the cost of implementation are high.
In order to keep flow attached to the volute walls and to keep the shape of the volute appropriate to the function of the volute, an A/R schedule is plotted, as in FIG. 5, to ensure that there exist no inappropriate changes in section. In FIG. 5, the “X” axis is the angle for each section. The angles could be substituted by the defining letters “A” though “M” as used in FIG. 4. The “Y” axis depicts the radius of the section. The dotted line (40) is the area schedule for the largest A/R of the family. The dashed line (41) is the area schedule for the smallest A/R of the family.
If one considers a wastegated turbo as a baseline for cost, then the cost of a typical (VTG) in the same production volume is from 270% to 300% the cost of the same size fixed, turbocharger. This disparity is due to a number of pertinent factors from the number of components, the materials of the components, the accuracy required in the manufacture and machining of the components, to the speed, accuracy, and repeatability of the actuator. The chart in FIG. 7 shows the comparative cost for the range of turbochargers from fixed to VTGs. Column “A” represents the benchmark cost of a fixed turbocharger for a given application. Column “B” represents the cost of a wastegated turbocharger for the same application, and column “C” represents the cost of a VTG for the same application.
Thus it can be seen that for both technical reasons and cost drivers that there needs to be a relatively low cost turbine flow control device which fits between wastegates and VTGs in terms of cost. The target cost price for such a device needs to be in the range of 145% to 165% that of a simple, fixed turbocharger.